KR940001323B1 - Variable compression ratio apparatus for internal combustion engine - Google Patents

Abstract

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Description

Variable compression ratio device of internal combustion engine

1 to 8 show a variable compression ratio device of an internal combustion engine as a first embodiment of the present invention.

1 is an overall configuration showing the appearance when in a low compression ratio.

2 is a front view of the connecting rod showing the shape when it is in a low compression ratio.

3 is an overall configuration showing the shape when the high compression ratio.

4 is a front view of the connecting rod showing the shape when it is in a high compression ratio.

5 is an enlarged cross-sectional view of the V portion of the first and third degrees.

6 is a sectional view of a hydraulic drive mechanism showing the shape when it is in a low compression ratio state.

7 is a sectional view of a hydraulic drive mechanism showing the shape when it is in a high compression ratio state.

8 is a hydraulic circuit diagram thereof.

9 to 13 show a variable compression ratio device of an internal combustion engine as a second embodiment of the present invention.

9 is an overall configuration diagram showing the shape when the low compression ratio state.

10 is an overall configuration showing the shape when in a high compression ratio state.

11 is a sectional view of a hydraulic drive mechanism showing the shape when it is in a low compression ratio state.

12 is a sectional view of a hydraulic drive mechanism showing the shape when it is in a high compression ratio state.

13 is a hydraulic circuit diagram thereof.

14 and 15 show a variable compression ratio device of an internal combustion engine as a third embodiment of the present invention.

14 is an overall configuration diagram in which the main part is broken and displayed.

Fig. 15 is a sectional view of the hydraulic drive mechanism.

16 and 17 show a variable compression ratio device of an internal combustion engine as a fourth embodiment of the present invention.

Figure 16 is an overall configuration diagram showing the main part broken.

17 is a sectional view of the hydraulic drive mechanism.

18 to 25 show a variable compression ratio device of an internal combustion engine as a fifth embodiment of the present invention.

18 is an overall configuration diagram showing the shape when the low compression ratio state.

19 is a front view of the connecting rod showing the shape when it is in a low compression ratio.

20 is an overall configuration showing the shape when in a high compression ratio state.

21 is an overall configuration diagram showing the shape when in a high compression ratio state.

22 is an enlarged cross-sectional view of part XXII of FIGS. 18 and 20;

23 is a sectional view of a hydraulic drive mechanism showing the shape when it is in a low compression ratio state.

24 is a cross-sectional view of the hydraulic drive mechanism showing the shape when in a high compression ratio state.

25 shows the hydraulic circuit thereof.

Fig. 26 is a view explaining the dimensions and the like of each part of the crankshaft.

27 is a characteristic diagram of hydraulic pressure relating to the hydraulic drive mechanism and its hydraulic supply system.

* Explanation of symbols for main parts of the drawings

(1): crankshaft (2): crankpin

(3): Crank journal (4): Crank arm

(5): eccentric sleeves (5a), (5b), (28a), (28b): engaging portions (notched)

(6): connecting rod (7): piston pin

(8): pistons (9), (10): metal bearings

(11): eccentric sleeve lock means

(11A), (35): Hydraulic drive mechanism (piston type hydraulic pressure drive mechanism)

(12), (30): Stopper pins (12a), (30a): Piston part

(13), (14), (31), (32): Hydraulic chamber (fluid pressure chamber)

(15), (33): literan spring (16), (34): cap

(17), (18): Hydraulic passage (fluid passage)

(19): oil pump (20): oil pan

(21), (26), (26 '): relief valve

(22): oil filter (23): main passage

(24): sub oil pump

(25), (25 '): switching valve (27), (27'): oil control valve

(28): eccentric sleeve (29): eccentric sleeve rock means

40: controller 41: engine load sensor

(42): engine speed sensor (51), (52): flange

(53): Notch part 54: Hydraulic cylinder

55: piston 56: pin

57: Window 58: Bushing Spring

(60), (61): orifice passage

The present invention relates to a variable compression ratio device of an internal combustion engine (hereinafter referred to as "engine" as necessary).

Conventionally, in the high load region or the engine rotation region larger than the engine heavy load region, knocking is prevented, while the compression ratio can be varied so as to improve fuel efficiency by improving the thermal efficiency in the operation region below the heavy load region. Various engines are proposed.

Such a compression ratio variable mechanism is disclosed, for example, in Japanese Patent Application Laid-Open No. 6332972. The compression non-variable mechanism disclosed in this publication has an eccentric bearing for eccentrically bearing a bearing hole of a connecting rod and an axis through which the bearing hole is inserted to one side of the burrs on both ends of the connecting rod of the engine. Hydraulic type to switch the rotation of the eccentric bearing between free and fixed by installing the lock pin to rotate freely by the rotational force generated by the reaction force from the support shaft and driving the lock pin which can move in the radial direction. An actuating lock means is provided, and the pressure of the supply working oil to the lock means is supplied to the lock means during the lock by a signal from the computer which receives a signal between the piston position detecting means and the operating condition detecting means. The pressure is applied and the control is performed under the condition that the hydraulic pressure is not applied to the locking means during unlocking.

However, in such a conventional variable compression ratio mechanism of an internal combustion engine, although the pressure of the supply operation oil to this lock means and the hydraulic force is always applied to the lock means during the lock, in this case, the operation oil is installed with a piston in between. Since only one side of the hydraulic chamber is operated, the hydraulic pressure in the hydraulic chamber varies greatly depending on the centrifugal force or the acceleration of the reciprocating motion of the connecting rod, which may cause the lock operation to be uncertain. There exists a problem that reliability falls.

The present invention has been made in view of the above problems, and provides a variable compression apparatus for an internal combustion engine, in which the locking and unlocking operation by the locking means is not uncertain in accordance with the centrifugal force or the acceleration of the reciprocating motion of the connecting rod. For the purpose of

An object of the present invention is to prevent a strong collision between the protrusion of the lock pin and the engagement portion of the eccentric bearing when switching the compression ratio, thereby preventing the protrusion of the lock pin or the engagement portion of the eccentric bearing with the lock pin. The present invention provides a variable compression ratio device for an internal combustion engine.

For this reason, the variable compression ratio apparatus of this invention is comprised as an apparatus containing the following characteristics.

(1) A connecting rod having one end supported by a piston reciprocating in the cylinder of the internal combustion engine and the other supported by a crankshaft at the other end thereof, and on either side of the support provided at both ends of the connecting rod. An eccentric sleeve which eccentrically aligns the bearing hole of the connecting rod and the support shaft through which the bearing hole is inserted therein is rotatably provided, and an eccentric sleeve lock means is formed to fix the rotation of the eccentric sleeve in a required position. The eccentric sleeve lock means is provided on both sides of the piston so that the pin member can engage with the engaging portion formed in the eccentric sleeve and the piston member is moved while moving the piston member connected to the pin member. Pieces having means for generating differential pressure between both fluid pressure chambers with fluid pressure required for the fluid pressure chambers respectively applied Provided with a hydraulic type drive mechanism is characterized in that configured.

(2) The piston fluid pressure drive mechanism (1) according to (1), wherein the piston-type fluid pressure drive mechanism includes a literal spring for biasing the piston portion toward the other fluid pressure chamber in one fluid pressure chamber, and is applied in advance to both fluid pressure chambers. It is possible to apply a fluid pressure higher than the required fluid pressure to the other fluid pressure chamber so that the means for applying the fluid pressure and the piston part to move toward the fluid pressure chamber against the side force of this literal spring. A means is provided.

(3) The method according to (2), comprising a first fluid pressure passage leading to the fluid pressure chamber and a second fluid pressure passage leading to the other fluid pressure chamber, and supplying a working fluid having a required pressure through the first fluid pressure passage. At the same time, a switching valve is provided in the second fluid passage, so that a working fluid having a required pressure or a working fluid having a pressure higher than this is supplied via the switching valve.

(4) The control valve (3) is characterized in that a control valve for switching control of the switching valve is provided while changing the pilot fluid pressure to the switching valve.

(5) The method according to (1), wherein the piston-type fluid pressure drive mechanism applies a fluid pressure to the both fluid pressure chambers so that a differential pressure is generated between the two fluid pressure chambers and the fluid pressure is also required at the low pressure side. And a means capable of switching between a state where the pressure side is high and a state where the fluid pressure side of the other fluid pressure chamber is high.

(6) The method of claim 5, wherein the first fluid passage leading to one fluid pressure chamber and the second fluid passage leading to the other fluid pressure chamber are provided, and a common switching valve is provided in the first fluid passage and the second fluid passage. And, through this switching valve, the first fluid passage and the second fluid passage are selectively supplied with a working fluid having a required pressure or a working fluid having a pressure higher than this.

(7) The control valve (8) is characterized in that a control valve for switching control of the switching valve is provided while changing the pilot fluid pressure to the switching valve.

⑧ The connecting rod is provided with a connecting rod which has one end supported by the piston reciprocating in the cylinder of the internal combustion engine and the other end supported by the crankshaft, and the nose supported part at the other end of the connecting rod. An eccentric sleeve for eccentrically aligning the bearing hole of the necking rod and the crankshaft through which the bearing hole is inserted is rotatably installed, and an eccentric sleeve lock means for fixing the rotation of the eccentric sleeve in the required position is provided. The eccentric sleeve lock means having a one-phase flange portion formed at both edges of the eccentric sleeve so as to sandwich the connecting rod therebetween, and a pin in which the eccentric sleeve lock member is protruded from the connecting rod toward both flange portions. A member and a flange portion of one side, and the pin member protrudes from the flange portion of the one side. And a second engaging portion provided at the flange portion on the other side and capable of engaging with the pin member when the pin member protrudes from the flange portion on the other side. Each of the piston chambers connected to the member is moved to the axial direction of the eccentric sleeve, so that the pin member can be driven into the fluid pressure chambers formed on both sides of the piston portion so as to be retractably driven from both of the connecting rods to both flange portions. And a piston type fluid pressure drive mechanism having a means for generating a differential pressure between both fluid pressure chambers in a state where a fluid pressure is applied.

(9) The through hole penetrating the connecting rod is formed in parallel with the bearing hole of the connecting rod so that the pin member is inserted into the through hole, and the central portion of the pin member is enlarged in diameter to form a piston. And a small diameter hole portion through which the through hole inserts one end of the pin member liquid tightly, an intermediate diameter hole portion accommodating the piston portion, and the other end portion of the pin member can be fluidly inserted. It consists of a three-stage hole part with a large diameter hole part to which a cap having a through hole therein is mounted, and two fluid pressure chambers are formed in the middle diameter hole part, so that this intermediate diameter hole part is configured as a cylinder part. It is characterized by.

The method of claim 8, wherein the first engagement portion and the second engagement portion are disposed out of phase by a required angle.

A metal bearing is installed between the eccentric sleeve and the eccentric sleeve between the bearing hole of the said connecting rod, and the metal bearing is retrofitted between the eccentric sleeve and the crankshaft.

The method of claim 1, wherein each of the metal bearings is formed with an endless fluid passage for supplying a working fluid to both fluid pressure chambers.

에 ⑧, wherein when the pin member engages the first engagement portion, the pin member enters a high compression ratio state or a low compression ratio state, and when the pin member engages with the second engagement portion, the pin member enters a low compression ratio state or a high compression ratio state. It is characterized by.

The fuel cell is characterized by a high compression ratio in the low engine speed medium region and a low compression ratio in the engine high speed zone.

절 ⑧, wherein the guide cutout portion cut in a circumferential direction in the circumferential direction on both circumferential surfaces of the flange portion, and the pin is guided to the guide cutout portion, the pin is a guide cutout And a shock absorbing mechanism for absorbing shocks abutting at both ends of the joint.

⑮항에 있어서, 상기 충격흡수기구가, 코넥팅로드내에 형성되어 양플랜지부에 대향하는 코넥팅로드 벽면에 긴구멍형상 창부가 형성된 실린더부와, 이 실린더부에 감삽되어 상기 창부를 통과하여 돌출하도록 상기 핀이 장착된 피스톤과, 상기 실린더부내에 있어서 이 피스톤의 양단을 각각 부세하는 스프링을 구비하여 구성된 것을 특징으로 한다.

The shock absorbing mechanism according to claim 1, wherein the impact absorbing mechanism is formed in a connecting rod and has a long hole-shaped window portion formed on a wall of the connecting rod facing the flange portions, and is inserted into the cylinder portion and protrudes through the window portion. It is characterized in that it is configured to include a piston to which the pin is mounted, and a spring that biases both ends of the piston in the cylinder portion.

항에 있어서, 상기 실린더부내에 있어서의 각 스프링 배설실에 통로를 개재해서 급배(給排)가능한 오일이 충진되어 있는 것을 특징으로 한다.

The oil supply that can be supplied and discharged through a passage is filled in each of the spring excretion chambers in the cylinder portion.

항에 있어서, 상기 각 스프링 배설실에 접속된 통로가 오리피스통로로서 구성된 것을 특징으로 한다.

A passage connected to each of said spring dropping chambers is configured as an orifice passage.

내연기관의 기통내를 왕복움직임하는 피스톤에 일단부를 돌쩌귀 지지되는 동시에 타단부를 크랭크샤아프트에 돌쩌귀 지지된 코넥팅로드를 구비하고, 이 코넥팅로드의 일단부에 있어서의 돌쩌귀 지지부에 상기 코넥팅로드의 베어링구멍과 이 베어링구멍을 삽통하는 피스톤핀을 상호 편심시키는 편심슬리이브가 회전가능하게 설치되고, 이 편심슬리이브의 회전을 소요위치에서 고정시킬 수 있는 편심슬리이브록수단이 설치되어서, 이 편심슬리이브록수단이, 상기 코넥팅로드내로부터 편심슬리이브의 축방향과 교차하는 방향으로 이동하므로서 이 편심슬리이브를 향해서 출몰가능하게 내장된 핀부재가 편심슬리이브에 상호 소요의 각도차를 가지고 설치되고 상기 핀부재가 편심슬리이브에 돌출해 나오면 각각 이 핀부재와 계합할 수 있는 제1, 제2의 계합부를 구비하는 동시에, 상기 핀부재에 연결된 피스톤부를 편심슬리이브의 축방향과 교차하는 방향으로 이동시키므로서 이 핀부재를 이 코넥팅로드내로부터 편심슬리이브를 향해서 출몰가능하게 구동할 수 있도록 이 피스톤부의 양쪽에 설치된 유체압실에 각각 소요의 유체압을 인가한 상태에서 양유체압실간에 차압(差壓)을 발생시키는 수단을 가진 피스톤식 유체압구동기구를 구비하여 구성된 것을 특징으로 한다.

A connecting rod which has one end supported by the piston reciprocating in the cylinder of the internal combustion engine and the other end supported by the crankshaft, and the connecting portion connected to the handle supporting part at one end of the connecting rod. An eccentric sleeve is provided rotatably to mutually eccentric the bearing hole of the rod and the piston pin through which the bearing hole is inserted, and an eccentric sleeve lock means is provided which can fix the rotation of the eccentric sleeve in a required position. As the eccentric sleeve lock means moves from the inside of the connecting rod in the direction intersecting with the axial direction of the eccentric sleeve, the pin member embedded in the eccentric sleeve has a angular difference of mutually required in the eccentric sleeve. And the first and second members, which are engaged with the pin member, when the pin member protrudes from the eccentric sleeve. In addition to having an engaging portion, it is possible to move the pin member connected to the pin member in a direction intersecting with the axial direction of the eccentric sleeve so that the pin member can be retractably driven from within the connecting rod toward the eccentric sleeve. And a piston-type fluid pressure drive mechanism having a means for generating a differential pressure between both fluid pressure chambers in a state in which required fluid pressure is applied to the fluid pressure chambers provided on both sides of the piston part.

항에 있어서, 상기 핀부재가 상기 제1계합부에 계합하면, 고압축비상태 또는 저압축비상태로되고, 상기 핀부재가 상기 제2계합부에 계합하면, 저압축비상태 또는 고압축비상태로 되는 것을 특징으로 한다.

The method of claim 8, wherein when the pin member engages in the first engagement portion, the pin member is in a high compression ratio state or a low compression ratio state, and when the pin member engages in the second engagement portion, it is in a low compression ratio state or a high compression ratio state. It features.

항에 있어서, 엔진저중속영역에서 고압축비상태로하고, 엔진고속영역에서 저압축비상태로 하는 것을 특징으로 한다.

A high compression ratio in the low engine speed medium region, and a low compression ratio in the engine high speed zone.

내연기관의 기통내를 왕복움직임하는 피스톤에 일단부를 돌쩌귀 지지되는 동시에 타단부를 크랭크샤아프트에 돌쩌귀 지지된 코넥팅로드를 구비하고, 이 코넥팅로드의 양단부에 있어서의 돌쩌귀 지지부의 어느 한쪽에 이 코넥팅로드의 베어링구멍과 이 베어링구멍을 삽통하는 지축을 상호 편심시키는 편심슬리이브가 회전가능하게 설치되고, 이 편심슬리이브의 축방향으로 이동할 수 있는 핀부재를 작동시켜서 이 편심슬리이브로 형성된 계합부에 이 핀부재를 계합시키므로서 이 편심슬리이브의 회전을 소요위치에서 고정시킬 수 있는 편심슬리이브록수단이 설치된 것을 특징으로 한다.

A connecting rod which is supported by the piston which reciprocates in the cylinder of the internal combustion engine and which is supported by the crankshaft at the other end is provided on either side of the connecting rod at both ends of the connecting rod. An eccentric sleeve is provided rotatably to mutually eccentric the bearing hole of the connecting rod and the support shaft through which the bearing hole is inserted, and the pin member movable in the axial direction of the eccentric sleeve is operated to form the eccentric sleeve. An eccentric sleeve lock means is provided which can fix the rotation of the eccentric sleeve in a required position by engaging the pin member with the engaging portion.

According to such a configuration, according to the apparatus of the present invention, even if the pressure difference between the two fluid pressure chambers does not change according to the change of the engine speed or the crank angle, the centrifugal force or the acceleration of the reciprocating motion of the connecting rod is thereby affected. This has the advantage that the pin member operates reliably.

In addition, there is an advantage that the pin member can be reliably operated even when the discharge capacity of the pressurizing means for supplying a high fluid pressure is low.

In addition, according to the present invention, since the shock generated between the stopper pin and the eccentric sleeve generated when the compression ratio is switched is absorbed by the hydraulic piston, the stopper pin and the part of the eccentric sleeve which abuts the stopper pin. It is possible to provide a variable compression ratio device for an internal combustion engine that can prevent the destruction of fuel in advance.

In addition, even when the inertia force is generated based on the reciprocating motion of the connecting rod, the operation of the pin member does not become uncertain, and there is an advantage of improving the reliability of the locking operation of the eccentric sleeve.

DETAILED DESCRIPTION Hereinafter, embodiments of the present invention will be described in detail with reference to the accompanying drawings.

First, the first embodiment will be described.

By the way, as shown in FIGS. 1-4, the piston pin of the piston 8 which makes the small end part reciprocate the inside of the cylinder of a gasoline engine (internal combustion engine) ( At the same time, the end of the crank piece 2 of the crankshaft 1 is supported.

Moreover, the eccentric sleeve 5 which mutually eccentrically aligns the bearing hole of the connecting rod 6 and the crank pin 2 as a support shaft through which this bearing hole is inserted into the support part at the large end of the connecting rod 6. ) Is rotatably installed. In other words, the eccentric sleeve 5 has an eccentric center of the inner circumferential circle and a center of the outer circumferential circle, and rotates the outer circumference of the crank pin 2 by 180 ^o from the maximum position of the eccentric to take the minimum eccentric position. It is.

In addition, between the inner circumferential surface of the eccentric sleeve 5 and the outer circumferential surface of the crank pin 2, as shown in detail in FIG. 5, the metal bearing 9 with the inner circumferential surface of the eccentric sleeve 5 is attached. The metal bearing 10 with the inner circumferential surface of the bearing hole of the connecting rod 6 is mounted between the outer circumferential surface of the eccentric sleeve 5 and the inner circumferential surface of the bearing hole of the connecting rod 6. . Accordingly, it is possible to slide between the eccentric sleeve 5 and the crank pin 2, and to slide between the eccentric sleeve 5 and the bearing hole of the connecting rod 6 at the same time. It is possible to slide between the eccentric sleeve 5 and the bearing hole of the connecting rod 6.

By the way, although the eccentric sleeve lock means 11 is provided, this eccentric sleeve lock means 11 is a pin member which is movable in the axial direction of the eccentric sleeve 5, ie, in the axial direction of the crankshaft 1. A stopper pin 12 is provided, and the stopper pin 12 is operated by the hydraulic drive mechanism 11A as the piston type hydraulic pressure drive mechanism, thereby providing two engagement portions 5a formed in the eccentric sleeve 5. ) And (5b) engage the stopper pins 12 so that the rotation of the eccentric sleeve 5 is at two positions (with the connecting rod 6 interposed at both edges of the above-mentioned single-sided maximum position). Each of the pair of flange portions formed and the minimum eccentric position).

This eccentric sleeve lock means 11 is described in detail. As shown in Figs. 6 and 7, first, the piston portion 12a is provided integrally with an enlarged diameter in a flange shape in the middle portion of the stopper pin 12, and the stopper pin with the piston portion 12a ( 12 is fitted into the through-hole formed in the large end of the connecting rod 6. The through hole penetrates the large end of the connecting rod 6 in the crankshaft axial direction, and is configured as a three-stage hole having three diameters, and the small diameter hole located at one end thereof has a stopper pin 12 The diameter of the middle portion is almost the same as that of the piston 12a, and the large diameter hole portion located at the other end is set larger than the diameter of the piston 12a.

Therefore, when the stopper pin 12 with the piston portion 12a is inserted into the through hole, the stopper pin 12 is inserted into the small diameter hole portion of the through hole in a liquid-tight manner, and the piston portion in the middle diameter hole portion of the through hole. 12a is inserted in a liquid tight manner. Then, the lean spring ring 15 is inserted, and a cap 16 having a through hole having a diameter substantially the same as that of the stopper pin 12 is formed in the center at a diameter substantially the same as that of the large diameter portion of the through hole. When the 16 is fixed to the connecting rod 6 with a bolt or the like, the stopper pin 12 is tightly inserted into the small diameter portion of the through hole while the other end thereof is liquid sealed into the through hole of the cap 16. To be subtracted, the middle diameter portion of the through hole is divided into two chambers 13 and 14 in the piston portion 12a. Inflow passages 17 and 18 are connected to the chambers 13 and 14, respectively. Accordingly, these two chambers are configured as hydraulic chambers (fluid pressure chambers) 13 and 14 formed on both sides of the piston portion 12a. In addition, the literal spring 15 is mounted in the hydraulic chamber 13 so that the stopper pin 12 with the piston portion 12a is attached to the hydraulic chamber 14 side. In addition, the hydraulic pressure area of both piston parts 12a is set equally.

Thereby, the piston part connected to the stopper pin 12 by the piston part 12a, the hydraulic chamber 13, 14, the liter spring 15, the cap 16, etc. which were attached to this stopper pin 12 was carried out. 11A of piston type hydraulic drive mechanisms which can drive the stripper pin 12 by moving 12a are comprised.

In addition, the eccentric sleeve 5 has a plan portion isolated in the axial direction with the large end of the connecting rod 6 interposed therebetween, as shown in the first to fourth cavities, but the eccentric sleeve in one flange portion. The notch-shaped engaging part 5a is formed in the part which the eve 5 takes an eccentric minimum position, and the notch is formed in the part which the eccentric sleeve 5 in the other flange part minimizes a eccentric maximum position. As shown in Figs. 3 and 4, in the state where the engaging portion 5b of the shape is formed and the stopper pin 12 is moved to the right and the first position is taken as shown in Figs. Likewise, the stopper pin 12 and the engaging portion 5b are engaged so that the eccentric sleeve 5 is fixed to the connecting rod 6 at the largest end in the maximum eccentric position, and the stuffer pin 12 is the first. , As shown in Fig. 6, while moving to the left and taking the second position, as shown in Figs. 1 and 2, Stopper pin 12 and the engaging portion (5a) is engaged by the eccentric sleeve (5) is adapted to be fixed to the big end of the connecting rod nose (6) in at least an eccentric position.

In addition, the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position, the connecting rod 6 is in the most elongated state in appearance, thereby realizing a high compression ratio. When the sleeve 5 is fixed to the large end of the connecting rod 6 in the minimum eccentric position, the connecting rod 6 is in a state that is most rounded in appearance, thereby realizing a low compression ratio state. In addition, the compression ratio in this low compression ratio state is selected so that the engine does not generate knocking, which is almost equal to the value set in the normal engine. Therefore, the compression ratio in the high compression ratio state is set to a value higher than the value set in the normal engine.

In addition, a means for applying hydraulic pressure (standard hydraulic pressure) required in advance through the hydraulic passages 17 and 18 to the hydraulic pressure chambers 13 and 14, and the stopper pin 12 with the piston portion 12a. Means for applying a hydraulic pressure (standard hydraulic pressure + alpha) higher than the above-mentioned standard hydraulic pressure to the hydraulic chamber 14 to move it toward the hydraulic chamber 13 against the side force of the liter spring 15. It is installed.

That is, the hydraulic passages 17 and 18 pass through the portion of the crank arm 4 from the crank journal 3 of the crankshaft 1 as shown in Figs. 2) passes through the large ends of the metal bearing 9, the eccentric sleeve 5, the metal bearing 10 and the connecting rod 6, and is connected to the hydraulic chambers 13 and 14, respectively. .

In addition, since the metal bearing 9 and the crank piece 2 and the metal bearing 10 and the eccentric sleeve 5 slide between them, as shown in FIG. 5, the metal bearing 9, On the inner circumferential surface of (10), two rows of endless grooves are formed around the inner circumferential surface and connected to the hydraulic passages (17) and (18), respectively, and each groove formed in the metal bearing (9) has an eccentric sleeve (5). The through holes are formed in the hydraulic passages (17) and (18) formed in the above, and the hydraulic passages formed in the large ends of the connecting rods (6) are also formed in the grooves formed in the metal bearing (10). Through-holes are formed in portions of (17) and (18).

As shown in FIG. 8, a portion other than the crankshaft of the hydraulic passage 17 is connected to the main passage 23 side, and a portion other than the crankshaft of the hydraulic passage 18 is connected to the sub oil pump 24. As shown in FIG. ) Or the main passage (23) side. That is, the oil (lubricating oil) from the oil tank or the oil pan 20 is the oil of the required hydraulic pressure (hydraulic supplying the standard hydraulic pressure) by the oil pump 19 with the relief valve 21 and the oil filter 22. Is supplied to the main passage 23 through the main passage 23, and the oil of standard hydraulic pressure is supplied through the hydraulic passage 17 from the main passage 23. The oil from the main passage 23 is supplied to the sub oil pump 24 and discharged again as a high oil pressure (standard oil pressure + alpha), but the oil pressure from the sub oil pump 24 is a switching valve ( The hydraulic pressure from the main passage 23 and the hydraulic passage 18 are selectively supplied via the passage 25. In other words, when the switching valve 25 is in the a position as shown in FIG. 8, the hydraulic passage 18 is supplied with the standard hydraulic pressure from the main passage 23, and the switching valve 25 is in the b position. The hydraulic passage 18 is supplied with a high hydraulic pressure (standard hydraulic pressure + alpha) from the sub oil pump 24.

Therefore, when the switching valve 25 is set to the b position, the hydraulic passage 18 is supplied with high hydraulic pressure (standard hydraulic pressure + alpha) from the sub-oil pump 24, and the high hydraulic pressure is supplied to the hydraulic chamber 14. do. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 13 via the hydraulic passage 17, the stopper pin with the piston part 12a is resisted by the force of the liter spring 15. If (12) moves to the right as shown in Figs. 3 and 7, and takes the first position, as shown in Figs. 3 and 4, the stopper pin 12 and the engaging portion 5b are engaged. Thus, the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 in the maximum eccentric position. As a result, the connecting rod 6 is in an elongated state in appearance, thereby realizing a high compression ratio.

When the switching valve 25 is set to the a position, the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic passage 18, and the standard hydraulic pressure is also supplied to the hydraulic chamber 14. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 13 via the hydraulic passage 17, the stopper pin with the piston portion 12a is applied by the force of the liter spring 15. When 12 moves to the left as shown in Figs. 2 and 6, and takes the second position, as shown in Figs. 1 and 2, the stopper pin 12 and the engaging portion 5a are engaged. Thus, the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 in the minimum eccentric position. As a result, the connecting rod 6 is in a state most depressed, so that a low compression ratio state can be realized.

The oil pump 19 and the sub oil pump 24 are driven together by an engine.

Reference numeral 26 in FIG. 8 denotes a relief valve, and the relief valve 26 adjusts the pressure difference α between the standard hydraulic pressure + alpha and the standard hydraulic pressure to be constant.

Further, reference numeral 27 denotes an oil control valve for switching control of the switching valve 25. When the oil control valve 27 is set to the a position, the pilot oil pressure of the switch valve 25 decreases and the switching valve 25 In the a position, and when the oil control valve 27 is in the b position, the pilot hydraulic pressure of the switching valve 25 is increased so that the switching valve 25 is in the b position.

Although the switching control signal from the controller 40 is input to the oil control valve 27, the controller 40 receives the detection signal from the engine load sensor 41 or the engine speed sensor 42. In response to the detection of the engine high load region or the engine high rotation region larger than the load region of the engine, the control signal for making the oil control valve 27 to the a position is outputted. When the area under the engine heavy load is detected, the oil control valve 27 Control signal is set to b position.

According to the above-described configuration, when the area below the engine heavy load is detected, the control signal for making the oil control valve 27 to the b position is output. Therefore, the switching valve 25 also becomes the b position, and the hydraulic passage 18 The high oil pressure from the oil pump 24 is supplied, and this high oil pressure (standard oil pressure + alpha) is supplied to the oil pressure chamber 14. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 13 via the hydraulic passage 17, as a result, the differential pressure α between the high hydraulic pressure (standard hydraulic pressure + α) and the standard hydraulic pressure is As a result, the differential pressure α decreases the negative force of the literian spring 15 and moves the stopper pin 12 with the piston portion 12a to the right as shown in FIGS. Move it. As a result, the stopper pin 12 takes the first position, and as shown in FIGS. 3 and 4, the stopper pin 12 and the engaging portion 5b engage, and the eccentric sleeve 5 is the maximum eccentricity. It is fixed at the end to the connecting rod 6 in position. As a result, the connecting rod 6 is in a state where it is most elongated in appearance, and is in a high compression ratio. In such a high compression ratio, the thermal efficiency is improved, and the fuel economy can be improved.

In addition, when the engine high load region or the engine high rotation region is detected that is larger than the engine heavy load region, a control signal for setting the oil control valve 27 to the a position is output. Therefore, the switching valve 25 also becomes the a position, so that the hydraulic passage ( 17) and 18 are supplied with standard hydraulic pressure from the main passage 23, and standard hydraulic pressure is supplied to the hydraulic chambers 13 and 14. As shown in FIG. Accordingly, when the stopper pin 12 with the piston portion 12a moves to the left as shown in the second and sixth degrees by the force of the literal spring 15 and takes the second position, As shown in Figs. 1 and 2, the stopper pin 12 and the engaging portion 5a are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. . As a result, the connecting rod 6 is in a state that is most depressed in appearance, resulting in a low compression ratio. Thus, knocking can be reliably avoided by setting it as a low compression ratio state.

Thus, in this first embodiment, since the control hydraulic pressure of the standard hydraulic pressure is applied to the hydraulic oil pressure chambers 13 and 14, respectively, as a base, the pressure difference between the hydraulic oil pressure chambers does not change due to the change of the engine speed or the crank angle. Therefore, the operation of the stopper pin 12 does not become uncertain due to the centrifugal force or the acceleration of the reciprocating motion of the connecting rod 6.

In addition, since the stopper pin 12 is configured to move in the axial direction of the crankshaft 1, the stopper pin 12 is reliably operated even if the stopper pin 12 is affected by the inertia force generated by the reciprocating motion of the connecting rod or the like. You can.

In addition, when the stopper pin 12 takes the first position, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the maximum eccentric position, and the stopper pin 12 is reversed. When the second position is moved in the direction, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the minimum eccentric position. Therefore, the drive timing of the stopper pin 12 needs to be considered. Therefore, the control for changing the compression ratio can be simplified. On the other hand, in the Japanese Unexamined Patent Application Publication No. 63-32972, it is necessary to control the lock pin protrusion timing by detecting the piston position, and the control becomes complicated.

As the control hydraulic pressure supplied to the hydraulic chambers 13 and 14 as described above is supplied through the path of the crank pin 2 → connecting rod 6 of the cylinder block journal → crankshaft 1, This control hydraulic pressure fluctuates greatly under the crankshaft centrifugal force or the connecting rod centrifugal force.

That is, as shown in FIG. 26, the crankshaft angular velocity is ω, the crank radius is R, the crank journal radius is rj, the crank pin radius is rp, and the flow path length from the center of the crank pin to the center of the stopper pin is l. When the hydraulic pressure of the journal portion is P _A , the hydraulic pressure P _PC at the center of the crank pin at the angular velocity ω is as follows.

P _PC = P _A + (1 / A) ∫ △ mrω ² = P _A + (1/2) ρω ² (R ² -rj ² ) ----- (1)

Here, A is the cross-sectional area of the flow path, Δm is the mass of the oil between the micro distance, ρ is the density of the oil.

In addition, the connecting rod supply hydraulic pressure P _PC becomes as follows.

P _CR = P _PC + (1 / A) ∫ △ mrω ²

_{= P A + (1/2) ρω} 2 (R 2 -rj 2) + ρω 2 · R · rp · cosωt ------ (2)

Similarly, the connecting rod stopper pin part hydraulic pressure P _SP becomes as follows.

P _SP = P _A + (1/2) ρω ² (R ² -rj ² ) + ρω ² R (rp + l) cosωt

= P _A + ρω ² [(1/2) (R ² -rj ² ) + R (rp + l) cosRωt ------ (3)

here,

P _B = (1/2) ρω ² (R ² -rj ² ) ---------------- (4)

PC = ρω ² (rp + l) cosωt ---------------- (5)

In this case, since P _A is a function of engine speed and oil viscosity coefficient, P _B is a function of engine speed and crank phase, the control hydraulic pressure of the stopper pin built in the connecting rod 6 is shown in FIG. It fluctuates greatly as shown. In Fig. 27, the connecting rod supply hydraulic pressure P _CP is the crank pin floating pressure (max, min) (characteristics B and C), and the connecting rod and stopper pin floating pressure P _SP is the stopper pin floating pressure (max, min). (Characteristics D and E).

However, if the differential pressure? Acts between the hydraulic oil pressure chambers 13 and 14 as in the present embodiment, the pressure fluctuating under the crankshaft centrifugal force or the connecting rod centrifugal force cancels out, so that the engine speed and the crank phase It is possible to form a hydraulic supply system that is not affected by.

Next, a second embodiment will be described using FIGS. 9 to 13.

By the way, this 2nd Example also makes it possible to insert the bearing hole and the bearing hole of the connecting rod 6 into the support part in the big end of the connecting rod 6, as shown to FIG. 9, 10 degree. The eccentric sleeve 5 which mutually eccentrically cranks the pin 2 as a support shaft is rotatably provided so as to take a maximum position and a minimum eccentric position at one time.

An eccentric sleeve lock means 11 is also provided, and the eccentric sleeve lock means 11 also moves in the axial direction of the eccentric sleeve 5, that is, in the axial direction of the crankshaft 1. The stopper pin 12 is operated by the hydraulic drive mechanism 11A as the piston-type fluid pressure drive mechanism, so that the stopper pin 12 is connected to the two engaging portions 5a and 5b formed on the eccentric sleeve 5. The rotation of the eccentric sleeve 5 can be fixed at two positions (the eccentric maximum position and the minimum eccentric position).

Moreover, also in this embodiment, the eccentric sleeve 5 has the flange column isolate | separated axially across the large end of the connecting rod 6, and the eccentric sleeve 5 in one flange part is carried out. A notch-shaped engaging portion 5a is formed at a portion where the eccentric minimum position is formed, and a notched system is formed at a portion where the eccentric sleeve 5 at the other flange portion has an eccentric maximum position. As shown in FIG. 10, a stopper pin (b) is formed and the stopper pin 12 moves to the right as shown in FIGS. 10 and 12 to take the first position. 12) and the engagement number 5b are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position, and the stopper pin 12 is shown in FIGS. 9 and 11 degrees. As shown in Fig. 9, the stopper is moved to the left and the second position is taken. The pin 12 and the engaging portion 5a are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position, and the eccentric sleeve 5 is the maximum eccentricity. When it is fixed to the large end of the connecting rod 6 in the position, the connecting rod 6 is in the most elongated state in appearance, thereby realizing a high compression ratio state, and the eccentric sleeve 5 is the nose in the minimum eccentric position. When the fixing rod 6 is fixed to the large end of the connecting rod 6, the connecting rod 6 is in a state where it is most elongated in appearance, so that a high compression ratio state can be realized.

Incidentally, in this second embodiment, the structure of the eccentric sleeve lock means 11 is different from that of the second embodiment described above.

That is, the stopper pin 12 with the piston portion 12a is fitted into the through hole formed in the large end of the connecting rod 6, and the cap 16 is fixed to the connecting rod 6 with a bolt or the like. The middle diameter portion of the through hole is divided into two chambers 13 and 14 in the piston portion 12a, which is configured as a hydraulic chamber, and the hydraulic passages 17 and 14 in the chambers 13 and 14, respectively. The connection of 18 is the same as that of the first embodiment described above, but in this second embodiment, the literal spring is not loaded in either of the hydraulic chambers 13 and 14.

Instead, the hydraulic pressure chambers 13 and 14 are formed so that the differential pressure α '(α' <α) is generated between the hydraulic oil pressure chambers 13 and 14 to have the hydraulic pressure (standard hydraulic pressure) required even at the low pressure side. ), And the hydraulic side of one hydraulic chamber 13 is high (standard hydraulic pressure + α ') and the hydraulic side of the other hydraulic chamber 14 is high (standard hydraulic pressure + α'). It has a means to switch.

This means will be described in detail again.

That is, the hydraulic passages 17 and 18 pass through the portion of the crank arm 4 from the crank journal 3 of the crankshaft 1, as shown in Figs. 2) passes through the large ends of the metal bearing 9, the eccentric sleeve 5, the metal bearing 10 and the connecting rod 6, and is connected to the hydraulic chambers 13 and 14, respectively. As shown in FIG. 13, portions other than the crankshafts of the hydraulic passages 17 and 18 are connected to the sub oil pump 24 or the main passage 23 side. That is, the oil (lubricating oil) from the oil pan 20 is the oil of the required hydraulic pressure (standard hydraulic pressure) by the oil pump 19 with the relief valve 21 via the oil filter 22 through the main passage 23. ), The oil from the main passage 23 is supplied to the sub oil pump 24 and discharged again as a high oil pressure (standard oil pressure + alpha '), but from the sub oil pump 24 The oil pressure is supplied to the oil pressure from the main passage 23 and the hydraulic passages 17 and 18 selectively via the switching valve 25 '. That is, when the switching valve 25 'is in the α position as shown in FIG. 13, the hydraulic passage 17 is supplied with the standard hydraulic pressure from the main passage 23, and the hydraulic passage 18 is sub-oiled. While the high hydraulic pressure (standard hydraulic pressure + alpha ') from the pump 24 is supplied and the switching valve 25 is set to the b position, the hydraulic pressure passage 18 is supplied with the standard hydraulic pressure from the main passage 23. The high pressure oil pressure (standard oil pressure + alpha ') from the sub-oil pump 23 is supplied to the oil pressure passage 17. As a result, the differential pressure α 'is always generated between the hydraulic oil pressure chambers 13 and 14.

Therefore, when the switching valve 25 'is in the α position, the hydraulic passage 17 is supplied with the standard hydraulic pressure from the main passage 23, while the hydraulic passage 18 has a high hydraulic pressure from the sub-oil pump 24. (Standard hydraulic pressure + alpha ') is supplied, the standard hydraulic pressure is supplied to the hydraulic chamber 13, and the hydraulic pressure chamber 14 is supplied with (standard hydraulic pressure + alpha'). Accordingly, the stopper pin 12 with the piston portion 12a moves to the right as shown in Figs. 10 and 12, taking the first position, and as shown in Fig. 10, the stopper pin 12 And the engaging portion 5b engage, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position. As a result, the connecting rod 6 is in an elongated state in appearance, thereby realizing a high compression ratio.

When the switching valve 25 'is set to the b position, the hydraulic passage 18 is supplied with the standard hydraulic pressure from the main passage 23, and the hydraulic passage 17 has a high hydraulic pressure from the sub-oil pump 24. Since (standard hydraulic pressure + α ') is supplied, the stopper pin 12 with the piston portion 12a moves to the left as shown in Figs. 9 and 11, taking the position of Fig. 2 and displaying it in Fig. 9 As described above, the stopper pin 12 and the engaging portion 5a are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. As a result, the connecting rod 6 is in a state most depressed, so that a low compression ratio state can be realized.

In addition, the hydraulic pressure areas of both piston parts 12a are equally set, and the sliding friction of the piston part 12a is negligibly small.

Reference numeral 26 'in Fig. 13 is a relief valve, and the relief valve 26' is adjusted so that the differential pressure? 'Between (standard hydraulic pressure + alpha') and standard hydraulic pressure is constant.

Reference numeral 27 'denotes an oil control valve for switching control of the switching valve 25'. When the oil control valve 27 'is set to the a position, the pilot hydraulic pressure of the switching valve 25' is lowered and the switching valve is reduced. (25 ') can be set to the a position, and the oil control valve (27') to the b position will increase the pilot oil pressure of the switching valve (25 ') and make the switching valve (25') to the b position. It is supposed to be.

Although the switching control signal from the controller 40 is input to the oil control valve 27 ', the controller 40 detects the detection signal from the engine load sensor 41 or the engine speed sensor 42. When the engine high load region or the engine high rotation region is detected that is larger than the load region of the engine, the control signal is issued to set the oil control valve 27 'to the b position. The control signal at the position a is set to 27 '.

Although the switching control signal is inputted from the controller 40 to the oil control valve 27 ', the controller 40 receives the detection signal from the engine load sensor 41 or the engine speed sensor 42. When the engine high load region or the engine high rotation region is detected that is larger than the load region in the engine, the control signal is issued to set the oil control valve 27 'to the b position, and when the area below the engine heavy load is detected, the oil control valve ( 27 ') to give a control signal to position a.

According to the above-described configuration, when an area below the engine heavy load is detected, a control signal for setting the oil control valve 27 'to the a position is output. Thus, the hydraulic valve 17 is set to the a position with the switching valve 25'. ) Is supplied with the standard hydraulic pressure from the main passage 23, while the hydraulic passage 18 is supplied with the high hydraulic pressure (standard hydraulic pressure + alpha ') from the sub-oil pump 24, and the hydraulic pressure in the hydraulic chamber 14 is supplied. As a result, the differential pressure α 'between the high hydraulic pressure (standard hydraulic pressure + α') and the standard hydraulic pressure is applied to the piston portion 12a. As a result, the differential pressure α 'portion is applied to the piston portion 12a. The attachment stopper pin 12 is moved to the right as shown in Figs. 10 and 12 to take the first position. As a result, as shown in FIG. 10, the stopper pin 12 and the engaging portion 5b are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the maximum eccentric position. . As a result, the connecting rod 6 is in a state where it is most elongated in appearance, and is in a high compression ratio. In such a high compression ratio, the thermal efficiency is improved, and the fuel economy can be improved.

When the engine high load region or the engine high rotation region is detected which is larger than the engine heavy load region, a control signal for setting the oil control valve 27 'to the b position is output, so that the switching valve 25' also becomes the b position. The hydraulic passage 18 is supplied with the standard hydraulic pressure from the may passage 23, and the hydraulic passage 17 is supplied with the high hydraulic pressure (standard hydraulic pressure + alpha ') from the sub oil pump 24. Since the hydraulic pressure in the hydraulic chamber 13 is in a high state, the stopper pin 12 with the piston portion 12a moves to the left as shown in FIGS. 9 and 11 to take the second position, and the ninth position. As shown in the figure, the stopper pin 12 and the engagement number 5a engage, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. As a result, the connecting rod 6 is in a state that is most depressed in appearance, and thus is in a low compression ratio state. Thus, knocking can be reliably avoided by setting it as the low compression ratio state.

As described above, in the second embodiment, the differential pressure? 'Is generated between the hydraulic pressure chambers 13 and 14 so that the hydraulic pressure chambers 13 and 14 have a hydraulic pressure (standard hydraulic pressure) required even at the low pressure side. The hydraulic pressure is applied and the hydraulic side of one hydraulic chamber 13 is high (standard hydraulic pressure + α ') and the hydraulic side of the other hydraulic chamber 14 is high (standard hydraulic pressure-α'). Since there is a means capable of switching to and from the engine, the pressure difference between the hydraulic chambers does not change due to the change of the engine speed or the crank angle, and accordingly, the stopper pin 12 ), The pressure difference α 'between the hydraulic pressure chambers 13 and 14 is the pressure difference α in the first embodiment described above (this pressure difference α is the value of the literal spring 15). It can be small (in theory, positive random) The stopper pin 12 can be reliably driven even when the discharge capacity of the sub-oil pump 24 is low, for example, at the time of low engine revolution.

In addition, since the stopper pin 12 is configured to move in the axial direction of the crankshaft 1, the stopper pin 12 is reliably operated even if the stopper pin 12 is affected by the inertia force generated by the reciprocating motion of the connecting rod or the like. You can.

In addition, when the stopper pin 12 takes the first position, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the most eccentric position, and the stopper pin 12 is reversed. When the second position is moved in the direction, the eccentric sleeve 5 can be fixed to the large end of the connecting rod 6 at the minimum eccentric position. Therefore, the drive timing of the stopper pin 12 needs to be considered. Therefore, the control for changing the compression ratio can be simplified.

Next, the third embodiment will be described using FIGS. 14 and 15 degrees.

By the way, in the third embodiment, as shown in Fig. 14, the bearing hole of the connecting rod 6 and the bearing hole are inserted into the support portion at the small end of the connecting rod 6. An eccentric sleeve 28 which mutually eccentric the piston pins 7 as the support shaft is rotatably provided so as to take the maximum and minimum eccentric positions of the eccentricity.

In addition, an eccentric sleeve lock means 29 is provided, and the eccentric sleeve lock means 29 has a stopper pin 30 as a pin member that can move in a direction orthogonal to the axial direction of the eccentric sleeve 28. The stopper pin 30 is engaged with the two engagement portions 28a and 28b formed in the eccentric sleeve 28 while operating by the hydraulic drive mechanism 35 as the piston type hydraulic pressure driving mechanism. The rotation of the eve 28 can be fixed at two positions (the eccentric maximum position and the minimum eccentric position above).

In the eccentric sleeve 28, a notch-shaped engaging portion 28a is formed at the portion where the eccentric sleeve 28 at the outer peripheral portion thereof has the eccentric minimum position, and the eccentric sleeve at the outer peripheral portion thereof. A cutout engaging portion 28b is formed at the portion at which the 28 takes an eccentric maximum position. When the stopper pin 30 protrudes, the stopper pin 30 and the engaging portion 28a or the engaging portion 28b are formed. ) Is supposed to engage.

Then, engagement with either the stopper pin 30 and the engaging portion 28a or 28b is performed by the following means. That is, as disclosed in Japanese Unexamined Patent Application Publication No. 63-32972, the number of teeth of the ring gear that rotates together with the crankshaft 1 is counted by electromagnetic pickup, and thus the crank position and the piston The position is detected and the timing which engages with either the engagement part 28a or 28b from this piston position is taken.

Accordingly, when the eccentric sleeve 28 is fixed to the small end of the connecting rod 6 in the maximum eccentric position or the minimum eccentric position, the connecting rod 6 is in the elongated state or the most shrunken state in appearance. Therefore, a high compression ratio state or a low compression ratio state can be realized.

In addition, the stopper pin 30 is formed by integrally forming a piston portion 30a (the hydraulic pressure areas of both of the piston portions 30a are equally set) in the middle portion, or a stopper pin with the piston portion 30a. (30) is opened in the inner circumferential surface of the bearing hole of the small end of the connecting rod (6) and inserted into a recess of approximately the same diameter as the piston portion (30a), and then the releasing spring (33) is extended to the cap (34). ), The hydraulic chambers (fluid pressure chambers 31 and 32 are formed in both of the piston portions 30a. The hydraulic passages 17 are formed in these hydraulic chambers 31 and 32, respectively). (18) are connected to the hydraulic passages (17) and (18) through the connecting rods (6), the metal bearings (9), and the crankshaft (1), as shown in FIG. The parts other than the crankshaft 1 are connected to the infant circuit system as shown in FIG.

In addition, a means for applying hydraulic pressure (standard hydraulic pressure) required in advance through the hydraulic passages 17 and 18 to the oil pressure chambers 31 and 32, and the stopper pin 30 with the piston portion 30a. Means for applying a hydraulic pressure higher than the standard hydraulic pressure (standard hydraulic pressure + alpha) to the hydraulic chamber 32 so as to move the c) against the side force of the liter spring 33 and move toward the hydraulic chamber 31. It is installed. And when said high oil pressure acts on the hydraulic chamber 32, the stopper pin 30 protrudes and engages with either the engagement part 28a or 28b.

In addition, although the switching control signal from the controller 40 is input to the oil control valve 27 shown in FIG. 8, the controller 40 is different from the first embodiment described above, and the engine load sensor ( 41) or the engine speed sensor 42 and the electronic pick-up (piston position sensor), the detection signal is sent to Roxy, and the control signal for the oil control valve 27 to the b position is outputted to the stopper pin 30. In the case of changing the engaged part, the control signal for setting the oil control valve 27 to the a position is output. The timing for outputting the control signal is also determined based on the signal from the piston position sensor.

With the above-described configuration, when the area under the engine heavy load is detected, a control signal for turning the oil control valve 27 to the a position is output, and the switching valve 25 is set to the a position. Then, the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic passages 17 and 18 together, and the standard hydraulic pressure is supplied to the hydraulic chambers 13 and 14. As a result, the stopper pin 30 with the piston portion 30a is drawn in by the force of the literal spring 15, and the engagement between the stopper pin 30 and the engaging portion is released. In this way, after the locked state is released, the control signal for setting the oil control valve 27 to the b position is output at the required timing. As a result, the switching valve 25 is also in the b position, and the hydraulic passage 18 is supplied with the high hydraulic pressure from the oil pump 24, so that the high hydraulic pressure (standard hydraulic pressure + alpha) is supplied to the hydraulic chamber 32. do. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 31 via the hydraulic passage 17, as a result, the differential pressure α between the high hydraulic pressure (standard hydraulic pressure + α) and the standard hydraulic pressure is The differential pressure? Is thus resisted by the bias force of the liter spring 33 so as to project the stopper pin 30 with the piston portion 30a. As a result, the stopper pin 30 and the engaging portion 28b which project the above control signal in relation to the output timing engage with each other so that the eccentric sleeve 28 of the connecting rod 6 is positioned at the maximum eccentric position. It is fixed to the small end. As a result, the connecting rod 6 is in a state where it is most elongated in appearance, and is in a high compression ratio. By setting it as such a high compression ratio, thermal efficiency improves and fuel efficiency improvement can be expected.

In addition, even when an engine high load region or an engine high rotation region larger than the engine heavy load region is detected, first, a control signal for setting the oil control valve 27 to a position is output, and the stopper pin 30 engages with the engaging portion. Releases the lock state, releases the locked state, and outputs a control signal for turning the oil control valve 27 to the b position at a separate timing, so that the switching valve 25 also becomes the b position, and the hydraulic passage 18 The high oil pressure from the oil pump 24 is supplied to this oil supply, and this high oil pressure (standard oil pressure + alpha) is supplied to the oil pressure chamber 32. As a result, as in the case described above, the stopper pin 30 with the piston portion 30a is projected against the bias force of the liter spring 33 at the differential pressure α. Therefore, in response to the control signal output timing described above, the protruding stopper pin 30 and the engaging portion 28a engage, and the small end of the connecting rod 6 at the minimum eccentric position of the eccentric sleeve 28. Is fixed to. As a result, the connecting rod 6 is in a state that is most depressed in appearance, and thus is in a low compression ratio state. Thus, the low compression ratio is achieved, so that knocking can be reliably avoided.

Thus, also in this third embodiment, the eccentric sleeve lock means 29 is disposed at the small end of the connecting rod 6, and the high compression ratio and low timing are achieved by the protruding action and the protruding timing of the stopper pin 30. The point of realizing the compression ratio state is different from that of the first embodiment, and since the standard hydraulic pressure is applied to both the hydraulic chambers 31 and 32 as a base, respectively, the engine speed and the crank angle are changed. The pressure difference between the two hydraulic chambers does not change, and accordingly, the protruding and pulling-in operation of the stopper pin 30 does not become uncertain due to the centrifugal force or the reciprocating motion of the connecting rod 6 due to the acceleration or the like.

Next, a fourth embodiment will be described using FIGS. 16 and 17 degrees. By the way, in the fourth embodiment, as shown in Figs. 16 and 17, in the 14 and 15 degrees, the literal spring 33 is removed and the hydraulic circuit system according to the second embodiment described above ( (See Fig. 13), the differential pressure pressure α '(α' &lt; α) is generated between the hydraulic pressure pressure chambers 31 and 32 so as to have a hydraulic pressure (standard hydraulic pressure) required even at the low pressure side. The hydraulic pressure is applied to 31 and 32, and the hydraulic side of one hydraulic chamber 31 is high (standard hydraulic pressure + α ') and the hydraulic side of the other hydraulic chamber 32 is high (standard hydraulic pressure). + α ') is enabled.

However, also in this case, the controller 40 which outputs the switching control signal to the oil control valve 27 'shown in FIG. 13 is different from the second embodiment described above, and the engine load sensor 41 or the engine rotation. Upon receiving the detection signals from the water sensor 42 and the electronic pick-up (piston position sensor), Roxy issues a control signal with the oil control valve 27 'at the a position and engages with the stopper pin 28. In the case of changing the engagement portion, a control signal for setting the oil control valve 27 'to the b position is output. The timing for outputting the control signal is also determined based on the signal from the piston position sensor.

With the above-described configuration, when the area under the engine heavy load is detected, a control signal is first given to the oil control valve 27 'at the b position. Then, the switching valve 25 'is also in the b position, and the hydraulic passage 18 is supplied with the standard hydraulic pressure from the main passage 23, while the hydraulic passage 17 is supplied with the high hydraulic pressure from the sub-oil pump 24. (Standard hydraulic pressure + alpha ') is supplied, so that the hydraulic side in the hydraulic chamber 31 is in a high state, so that the stopper pin 30 with the piston portion 30a enters, and the stopper pin 30 and the engaging portion The engagement is released. In this manner, after the locked state is released, the control signal for turning the oil control valve 27 'to the a position is performed at the required timing. As a result, the switching valve 25 'is also in the a position, and the hydraulic passage 17 is supplied with the standard hydraulic pressure from the main passage 23, while the hydraulic passage 18 is supplied from the sub-oil pump 24. Since high oil pressure (standard hydraulic pressure + alpha ') is supplied and the hydraulic pressure in the hydraulic chamber 32 is in a high state, as a result, the differential pressure α' between the high oil pressure (standard hydraulic pressure + alpha ') and standard hydraulic pressure is the piston. The part 30a is caught, and accordingly, this differential pressure? 'Protrudes the stopper pin 30 with the piston part 30a. As a result, in relation to the control signal output timing described above, the stopper pin 30 and the engaging portion 28b engage, so that the eccentric sleeve 28 is at the small end of the connecting rod 6 at the maximum eccentric position. Is fixed to. As a result, the connecting rod 6 is in a state where it is most elongated in appearance, and is in a high compression ratio. In this manner, when the high compression ratio is achieved, the thermal efficiency is improved, and the fuel economy can be improved.

Also, in the case where an engine high load region or an engine high rotation region larger than the engine heavy load region is detected, first, a control signal for setting the oil control valve 27 'to the b position is output, and the stopper pin 30 and the engaging portion After releasing the lock and releasing the locked state, the control signal outputting the oil control valve 27 'to the a position is output at a separate timing, so that the switching valve 25' is also in the a position and the hydraulic passage Standard hydraulic pressure from the main passage 23 is supplied to the 17, and high hydraulic pressure (standard hydraulic pressure + alpha ') from the sub-oil pump 24 is supplied to the hydraulic passage 18, so that the hydraulic chamber 32 is supplied. As the oil pressure side inside becomes high, as a result, the differential pressure α 'between the high hydraulic pressure (standard hydraulic pressure + α') and the standard hydraulic pressure is applied to the piston portion 30a, and this differential pressure α 'portion is thus applied to the piston portion 30a. ) The attachment stopper pin 30 protrudes. As a result, the stopper pin 30 and the engaging portion 28a engage the above-mentioned control signal with the output timing so that the eccentric sleeve 28 is at the small end of the connecting rod 6 at the minimum eccentric position. Is fixed to. As a result, the connecting rod 6 is in a state that is most depressed in appearance, and thus is in a low compression ratio state. Thus, the low compression ratio is achieved, so that knocking can be reliably avoided.

In this way, in the same manner as in the second embodiment described above, the hydraulic pressure chamber 31 is formed such that the differential pressure α 'is generated between the hydraulic pressure chambers 31 and 32 so as to have a hydraulic pressure (standard hydraulic pressure) required even at the low pressure side. , Hydraulic pressure is applied to (32), and the hydraulic side of one hydraulic chamber 31 is high (standard hydraulic pressure + α ') and the hydraulic side of the other hydraulic chamber 32 is high (standard hydraulic pressure + α). Since there is a means capable of switching with '), the pressure difference between the hydraulic oil pressure chambers does not change with the change of the engine speed or the crank angle. In addition to the fact that the operation of (12) does not become uncertain, the differential pressure α 'between the oil pressure chambers 31 and 32 is the differential pressure α in the embodiment shown in FIGS. 14 and 15 described above (this differential pressure α Is smaller than the literal spring (15) to overcome the Therefore, the stopper pin 30 can be reliably derived even when the discharge capacity of the sub-oil pump 24 is low, for example, at the time of low engine revolution. have

Next, a fifth embodiment will be described. In this fifth embodiment, first, as shown in Figs. 18 to 21, the connecting rod 6 reciprocates its small end in the cylinder of the gasoline engine (internal combustion engine). It is rotatably supported by the piston pin 7 of the crank pin, and its big end is supported by the crank pin 2 of the crankshaft 1.

Moreover, the eccentric sleeve 5 which mutually eccentrically cranks the pinc 2 as a bearing axis of the connecting rod 6 and the bearing hole of the connecting rod 6 is inserted into the support at the end of the connecting rod 6. ) Is rotatably installed.

In other words, the center of the inner circumferential surface and the center of the outer circumferential surface of the eccentric sleeve 5 are eccentric. When the eccentric sleeve 5 is rotated about 160 ^o from the minimum eccentric position, the outer periphery of the crank pin 2 is rotated. It is possible to take near the maximum eccentric position.

Further, as shown in detail in FIG. 22 between the inner circumferential surface of the eccentric sleeve 5 and the outer circumferential surface of the crank pin 2, the metal bearing 9 with the inner circumferential surface of the eccentric sleeve 5 is refurbished. In addition, between the outer circumferential surface of the eccentric sleeve 5 and the inner circumferential surface of the bearing hole of the connecting rod 6, the metal bearing 10 with the inner circumferential surface of the bearing hole of the connecting rod 6 is refurbished. . As a result, the eccentric sleeve 5 and the crank pin 2 are slid, and at the same time, the eccentric sleeve 5 and the connecting hole 6 can be slid between the bearing holes.

By the way, although the eccentric sleeve lock means 11 is provided, this eccentric sleeve lock means 11 is a stopper as a pin member which is movable in the axial direction of the eccentric sleeve 5, that is, in the axial direction of the crankshaft 1. Two engagement portions 5a formed in the eccentric sleeve 5 are provided with a pin 12, and the stopper pin 12 is operated by the hydraulic drive mechanism 11A as the piston type fluid pressure drive mechanism. By engaging the stopper pin 12 with 5b, the rotation of the eccentric sleeve 5 can be fixed at two positions (near the minimum and maximum eccentric positions described above).

This eccentric sleeve lock means 11 is described in detail. As shown in FIG. 23 and FIG. 24, first, the piston part 12a is enlarged in a flange shape, and is integrally provided in the intermediate part of the stopper pin 12, and this piston part 12a is attached. The stopper pin 12 is fitted into the through hole formed in the large end of the connecting rod 6. The through hole penetrates the large end of the connecting rod 6 in the crankshaft axial direction, and is configured as a three-stage hole having three diameters, and the small diameter hole located at one end thereof has a stopper pin 12 The diameter of the middle portion is almost the same as that of the piston 12a, and the diameter of the middle diameter portion of the piston 12a is substantially the same as the diameter of the piston 12a.

Therefore, when the stopper pin 12 with the piston portion 12a is inserted into the through hole, the stopper pin 12 is inserted into the small diameter hole portion of the through hole in a liquid-tight manner, and the piston portion in the middle diameter hole portion of the through hole. 12a is inserted in a liquid tight manner. Then, a liter spring 15 is inserted, and a cap 16 having a through hole having a diameter substantially equal to that of the large diameter portion of the through hole is inserted, and the cap 16 is bolted to the connecting rod 6 with a bolt or the like. When it is fixed, the stopper pin 12 is tightly inserted into the small diameter portion of the through hole while the other end thereof is tightly inserted into the through hole of the cap 16 so that the middle diameter portion of the through hole is the piston portion 12a. ) Are divided into two chambers (13), (14). The hydraulic passages 17 and 18 communicate with the chambers 13 and 14, respectively. Accordingly, these two chambers are configured as hydraulic chambers (fluid pressure chambers) 13 and 14 formed at both ends of the piston portion 12a. Further, the literal spring 15 is mounted in the hydraulic chamber 13 so that the stopper pin 12 with the piston portion 12a is attached to the hydraulic chamber 14 side. In addition, the hydraulic pressure areas of both piston parts 12a are equally set.

Accordingly, the piston connected to the stopper pin 12 by the piston part 12a, the hydraulic chamber 13, 14, the liter spring 15, the cap 16, etc. which are attached to this stopper pin 12. 11A of piston type hydraulic drive mechanisms which can drive the stopper pin 12 by moving the part 12a are comprised.

Also, as shown in the eccentric sleeve 5 and FIGS. 18 to 21, the flanges 51 and 52 are axially isolated from each other with a large end of the connecting rod 6 interposed therebetween. However, at the portion where the eccentric sleeve 5 in the one flange portion 51 takes the minimum eccentric position, a notch engaging portion 5a is formed, and in the other flange portion 52, In the portion where the eccentric sleeve 5 takes the vicinity of the maximum eccentric position, a notched engaging portion 5b is formed.

And the notch 53 for a guide is formed in the position where the said flange parts 51 and 52 match with each other. In addition, a cylindrical hydraulic cylinder 54 is attached to the bottom of the large end of the connecting rod 6. In this hydraulic cylinder 54, the piston 55 is liquid-extracted, and the pin 56 protrudes from the window 57 provided in both sides of the hydraulic piston 55 from both sides of the piston 55, have. Between the end faces of the piston 55 and the side wall of the hydraulic cylinder 54, the bushing springs 58 and 59 are charged and filled with oil. In addition, both seals in the hydraulic cylinder 54 divided by the piston 55 are connected to the oil supply passage via the orifice passages 60 and 61. The piston 55 is held in a neutral position by the subordinate forces of the bushing springs 58 and 59.

Then, as shown in FIGS. 20 and 21, the stopper pin 12 is shown in the state where the stopper pin 12 moves to the right and takes the first position as shown in FIGS. 20 and 24. ) And the engaging portion 5b are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 near the maximum eccentric position and the stopper pin 12 is in FIGS. 18 and 23. As shown in FIG. 18 and FIG. 19, the stopper pin 12 and the engaging portion 5a engage with each other, and the eccentric sleeve 5 ) Is fixed to the large end of the connecting rod 6 in the minimum eccentric position.

Then, when the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 near the maximum eccentric position, the connecting rod 6 is in an elongated state in appearance, thereby realizing a high compression ratio. When the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position, the connecting rod 6 is in the most undulating state in appearance, thereby achieving a low compression ratio. In addition, the compression ratio in this low compression ratio state is selected such that the engine does not cause knocking, which is almost the result of being set in a normal engine. Therefore, the compression ratio in the high compression ratio state is set to a value higher than the value set in the normal engine.

In addition, a means for applying hydraulic pressure (standard hydraulic pressure) required in advance through the hydraulic passages 17 and 18 to the hydraulic pressure chambers 13 and 14, and a stopper with a piston 12a is applied to the liter spring. Means for applying a hydraulic pressure (standard hydraulic pressure + alpha) higher than the above-mentioned standard hydraulic pressure are provided in the hydraulic chamber 14 so as to move toward the hydraulic chamber 13 in response to the negative force of (15).

That is, the hydraulic passages 17 and 18 pass through the portion of the crank arm 4 from the crank journal 3 of the crankshaft 1 as shown in FIGS. 18 and 20. The pins 2 pass through the large ends of the methyl bearing 9, the eccentric sleeve 5, the metal bearing 10 and the connecting rod 6, and communicate with the hydraulic chambers 13 and 14, respectively. It is.

Further, since the metal bearing 9 and the crank pin 2 and the metal bearing 10 and the eccentric sleeve 5 are slid, as shown in FIG. 23, the metal bearing 9, On the inner circumferential surface of (10), two rows of endless grooves connected to the hydraulic passages (17) and (18) circumscribing the inner circumferential surface are formed, and each groove formed in the metal bearing (9) has an eccentric sleeve (5). The through-holes are formed in the parts of the hydraulic passages 17 and 18 formed at the same time, and the hydraulic passages are formed in the grooves formed in the metal bearings 10 at the ends of the connecting rods 6. The through-holes which are matched with the part of 17) and 18 are formed.

As shown in FIG. 25, a portion other than the crankshaft of the hydraulic passage 17 is connected to the main passage 23 side, and a portion other than the crankshaft of the hydraulic passage 18 is connected to the sub oil pump ( 24) or the main passage (23) side. That is, the oil (lubricating oil) from the oil tank or the oil pan 20 is interposed with the oil filter 22 as oil of required oil pressure (hydraulic supplying standard oil pressure) by the oil pump 19 with a relief valve. The main passage 23 is supplied, and the main passage 23 passes through the hydraulic passage 17 to supply oil of standard hydraulic pressure. The oil from the main passage 23 is supplied to the sub oil pump 24 and discharged again as a high oil pressure (standard oil pressure + alpha). However, the oil pressure from the sub oil pump 24 is a switching valve ( The hydraulic pressure from the main passage 23 is selectively supplied to the hydraulic passage 18 via 25. In other words, when the switching valve 25 is in the a position as shown in FIG. 25, the hydraulic passage 18 is supplied with the standard hydraulic pressure from the main passage 23, and the switching valve 25 is in the b position. The hydraulic passage 18 is supplied with a high hydraulic pressure (standard hydraulic pressure + alpha) from the sub oil pump 24. In other words, when the switching valve 25 is in the a position as shown in FIG. 25, the hydraulic passage 18 is supplied with the standard hydraulic pressure from the main passage 23, and the switching valve 25 is in the b position. The inflow passage 18 is supplied with a high oil pressure (standard oil pressure + alpha) from the sub oil pump 24.

Therefore, when the switching valve 25 is set to the b position, the high oil pressure (standard oil pressure + alpha) from the sub oil pump 24 is supplied to the oil pressure passage 18 so that the high oil pressure is supplied to the oil pressure chamber 14. do. At this time, this high hydraulic pressure is supplied to the hydraulic chamber 13. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 13 via the hydraulic passage 17, the stopper pin with the piston part 12a is resisted by the force of the liter spring 15. When 12 moves to the right as shown in Figs. 20 and 24 and takes the first position, as shown in Figs. 20 and 21, the stopper pin 12 and the engaging portion are shown. (B) engages, and the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 near the maximum eccentric position. As a result, the connecting rod 6 is in an elongated state in appearance, thereby realizing a high compression ratio.

When the switching valve 25 is set to the a position, the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic passage 18, and the standard hydraulic pressure is supplied to the hydraulic chamber 14. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 13 via the hydraulic passage 17, the stopper pin with the piston portion 12a is applied by the force of the liter spring 15. As shown in FIG. 18 and FIG. 23, when moving to the left and taking a 2nd position, as shown in FIG. 18 and FIG. 19, the stopper pin 12 and the engaging part 5a are shown. In this manner, the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. As a result, the connecting rod 6 is in a state most depressed, so that a low compression ratio state can be realized.

The oil pump 19 and the sub oil pump 24 are driven by an engine.

Reference numeral 26 in Fig. 25 denotes a relief valve, and the relief valve 26 is adjusted so that the differential pressure α between (standard hydraulic pressure-alpha) and the standard hydraulic pressure becomes constant.

Further, reference numeral 27 denotes an oil control valve for switching control of the switching valve 26. When the oil control valve 27 is set to the a position, the pilot oil pressure of the switching valve 25 decreases and the switching valve 25 Can be set to the a position, and when the oil control valve 27 is set to the b position, the pilot hydraulic pressure of the switching valve 25 is increased so that the switching valve 25 can be set to the b position.

Although the switching control signal from the controller 40 is input to the oil control valve 27, the controller 40 receives the detection signal from the engine load sensor 41 or the engine speed sensor 42. When the engine high load region or the engine high rotation region is detected that is larger than the engine increase load region, the control signal is set to the position A of the oil control valve 27, and when the area under the engine heavy load is detected, the oil control valve 27 Control signal is set to b position.

According to the above configuration, when the area under the engine heavy load is detected, the control signal for setting the oil control valve 27 to the b position is output, so that the switching valve 25 also becomes the b position, and the hydraulic passage 18 The high oil pressure from the oil pump 24 is supplied, and this high oil pressure (standard oil pressure + alpha) is supplied to the oil pressure chamber 14. At this time, since the standard hydraulic pressure from the main passage 23 is supplied to the hydraulic chamber 13 via the hydraulic passage 17, as a result, the differential pressure α between the high hydraulic pressure (standard hydraulic pressure + α) and the standard hydraulic pressure is The pressure difference is applied to the piston portion 12a, so that the differential pressure resists the negative force of the liter spring 15 so that the stopper pin 12 with the piston portion 12a is shown on the right as shown in FIGS. 20 and 24. Move to. As a result, the stopper pin 12 takes the first position, and as shown in FIGS. 20 and 21, the stopper pin 12 and the engaging portion 5b engage with each other, so that the eccentric sleeve 5 Is fixed to the large end of the connecting rod 6 near the maximum eccentric position. At this time, when the stopper pin 12 and the engaging portion 5b are engaged, the pin 56 abuts on the end of the guide cutout 53 and the shock is buffered by the oil damper function. Then, the connecting rod 6 is switched without being impacted by the stopper pin 12 in a state where it is almost elongated in appearance, thereby becoming a high compression ratio. In such a high compression ratio, the thermal efficiency is improved, and the fuel economy can be improved.

In addition, when the engine high load region or the engine high rotation region that is larger than the engine main load region is detected, the control signal for the oil control valve 27 to the a position is output. Therefore, the switch valve 25 also becomes the a position and the hydraulic passage ( 17) and 18 are supplied with standard hydraulic pressure from the main passage 23 together, and the standard hydraulic pressure is supplied to the hydraulic chambers 13 and 14. Accordingly, when the stopper pin with the piston portion 12a moves to the left side as shown in Figs. 19 and 23 by the force of the literal spring 15, the second position is taken. As shown in FIG. 19, the stopper pin 12 and the engaging portion 5a are engaged so that the eccentric sleeve 5 is fixed to the large end of the connecting rod 6 at the minimum eccentric position. At this time, when the stopper pin 12 and the engaging portion 5b engage, the pin 56 abuts against the end of the guide cutout 53, and the shock is buffered by the oil damper function. As a result, the connecting rod 6 is switched without being impacted by the stopper pin 12 in a state most shrunk in appearance, and is brought into a low compression ratio state. Thus, the low compression ratio is achieved, so that knocking can be reliably avoided.

In each of the above embodiments, pressure oil was used as the driving means of the stopper pins 12 and 30 of the eccentric sleeve lock means 11 and 29. Liquid or gas) may be used.

As the mechanism for driving the stopper pin 12, in addition to using the above-mentioned fluid pressure drive mechanism, a drive mechanism using an electromagnetic principle (for example, driving the stopper pin 12 using electromagnetic force) is used. ) Can also be used.

Claims (22)

The connecting rod (6) is provided with a connecting rod (6) which is supported by the piston (8) for reciprocating the cylinder of the internal combustion engine and the other end is supported by the crank shaft (1). Eccentric sleeves (5) and (28) for eccentrically bearing the bearing holes of the connecting rod (6) and the support shaft through which the bearing holes are inserted are provided at either end of the support portion at both ends thereof, Eccentric sleeve lock means (11) and (29) are formed to fix the rotation of the eccentric sleeves (5) and (28) at the required positions, and the eccentric sleeve lock means (11) and (29) Engaging parts 5a, 5b formed in the eccentric sleeves 5, 28: pin members 12, 30 that can engage the 28a, 28b and the pin member 12 And fluids provided at both of the piston parts 12a and 30a so as to drive the pin members 12 and 30 by moving the piston parts 12a and 30a connected to the valve 30. Pressure chamber 13), (14): Piston-type fluid pressure drive mechanisms (11A), (35) having means for generating a differential pressure between both fluid pressure chambers with fluid pressure required for (31), (32) applied, respectively. Variable compression device of an internal combustion engine, characterized in that provided with. The fluid type pressure chamber of claim 1, wherein the piston type fluid pressure drive mechanisms 11A and 35 move the piston portions 12a and 30a to one of the fluid pressure chambers 13 and 31. The literal springs 15 and 33 which are urged to move toward (14) and (32) are provided, and are required in advance in both fluid pressure chambers 13, 14, 31 and 32. Means for applying a fluid pressure to be applied, and the piston parts (12a), (30a) to the side of the hydraulic pressure chamber (13), (31) against the side force of the liter spring (15), (33) And a means for applying a fluid pressure higher than the required fluid pressure to the other fluid pressure chambers (14) and (32) so as to move. 3. The second fluid passage (18) according to claim 2, wherein the first fluid passage (17) leading to one of the fluid pressure chambers (13) and (31) and the second fluid passage (18) leading to the other fluid pressure chambers (14) and (32) are provided. And a working fluid of required pressure is supplied through the first fluid passage 17, and a switching valve 25 is provided in the second fluid passage 18, via the switching valve 25. A variable compression device for an internal combustion engine, characterized by being configured to supply a working fluid having a required pressure or a working fluid having a higher pressure than this. 4. A variable compression apparatus for an internal combustion engine according to claim 3, wherein a control valve (27) for switching control of said well switching valve (25) is provided while changing the pilot fluid pressure to said switching valve (25). 2. The piston-type fluid pressure drive mechanisms 11A and 35 have a differential pressure generated between the two fluid pressure chambers 13, 14: 31, and 32, and are required even at the low pressure side. The fluid pressure is applied to both fluid pressure chambers so as to have a fluid pressure, and the fluid pressure side of one of the fluid pressure chambers 13 and 31 is high and the fluid pressure side of the other fluid pressure chambers 14 and 32 is high. And a means capable of switching to a state. 6. The fluid passage of claim 5, wherein the first fluid passage (17) through one of the fluid pressure chambers (13), (31) and the second fluid passage (18) through the other fluid pressure chambers (14), (32) are connected. And a common switching valve 25 'is provided in the first fluid passage 17 and the second fluid passage 18, and the first fluid passage 17 is provided through the switching valve 25'. And a working fluid of a required pressure or a working fluid of a pressure higher than this is supplied to the second fluid passage (18). 7. The variable internal combustion engine according to claim 6, characterized in that a control valve (27 ') for switching control of the switching valve (25') is provided while changing the pilot fluid pressure to the switching valve (25 '). Compression device. The connecting rod (6) is provided with a connecting rod (6) which is supported by the piston (8) for reciprocating the cylinder of the internal combustion engine and the other end is supported by the crank shaft (1). The eccentric sleeve 5 which eccentrically aligns the bearing hole of the connecting rod 6 and the crankshaft 1 through which the bearing hole is inserted is provided at the stopper support part at the other end thereof. An eccentric sleeve lock means 11 is provided which can fix the rotation of the eccentric sleeve 5 at a required position. The eccentric sleeve lock means 11 is provided at both edge portions of the eccentric sleeve 5. A pair of flange portions formed with the connecting rod 6 interposed therebetween, a pin member 12 embedded in the connecting rod 6 toward both flange portions, and a flange portion on one side thereof. The pin member 12 is a flange of one side Protrudes into the first engagement portion 5a, which can engage with the pin member 12, and the flange portion of the other side, and the pin member 12 protrudes from the flange portion of the other side. The second engaging portion 5b that can engage with the pin member 12 and at the same time moves the piston portion 12a connected to the pin member 12 in the axial direction of the eccentric sleeve 5. In each of the fluid pressure chambers 13 and 14 formed on both sides of the piston portion 12a, the pin member 12 can be driven in and out of the connecting rod 6 toward both flange portions. A variable compression apparatus for an internal combustion engine, comprising a piston-type fluid pressure drive mechanism 11A having a means for generating a differential pressure between both the fluid pressure chambers 13 and 14 in a state where required fluid pressure is applied. . The connecting rod 6 is formed with a through hole penetrating the connecting rod 6 in parallel with a bearing hole, and the pin member 12 is inserted into the through hole. A small diameter bore portion in which the central portion of the pin member 12 is enlarged in diameter to form the piston portion 12a, and the through-hole inserts one end of the pin member 12 in a liquid tight manner; It is configured as a three-stage hole portion having a middle diameter hole portion for accommodating (12a) and a large diameter hole portion on which a cap 16 having a through hole capable of fluidly inserting the other end portion of the pin member 12 is mounted. And at the same time, two fluid pressure chambers (13) and (14) are formed in the middle diameter hole, so that the middle diameter hole part is configured as a cylinder part. 9. The variable compression apparatus for an internal combustion engine according to claim 8, wherein the first engagement portion (5a) and the second engagement portion (5b) are disposed out of phase by a required angle. 9. The metal bearing 10 is mounted between the bearing hole of the connecting rod 6 and the eccentric sleeve 5, and at the same time, the eccentric sleeve 5 and the crankshaft 1 are separated. A variable compression device for an internal combustion engine, characterized in that a metal bearing (9) is interposed therebetween. 12. The internal combustion according to claim 11, wherein the metal bearings (9) and (10) are formed in an endless fluid passage for supplying a working fluid to both the fluid pressure chambers (13) and (14). Variable compression system of the engine. The method according to claim 8, wherein when the pin member 12 engages the first engagement portion 5a, the pin member 12 is in a high compression ratio state or a low compression ratio state, and the pin member 12 is in the second engagement portion 5b. And a low compression ratio state or a high compression ratio state. 14. The variable compression apparatus for an internal combustion engine according to claim 13, wherein the high compression ratio is set in an engine low medium speed region and a low compression ratio is set in an engine high speed region. The guide cutout 53 and the guide cutout 53 which are cut out by a predetermined section in the circumferential direction on both side circumferential surfaces of the flanges 51 and 52, according to claim 8; A variable compression device for an internal combustion engine, characterized by having a pin (56), the pin (56) having a shock absorbing mechanism for absorbing shocks abutting on both ends of the guide cutout (53). 16. The cylinder portion according to claim 15, wherein the shock absorbing mechanism is formed in the connecting rod (6) and has a long hole-shaped window portion (57) formed on the wall of the connecting rod (51) facing the two flange portions (51) and (52). (54), a piston (55) on which the pin (56) is mounted so as to be subtracted from the cylinder portion (54) and protrudes through the window portion (57), and both ends of the piston (55) in the cylinder portion. A variable compression device for an internal combustion engine, characterized in that it is provided with springs (58) and (59) respectively biasing. 17. The variable compression apparatus for an internal combustion engine according to claim 16, wherein oil is expelled and filled through each passage in the spring discharge chamber in the cylinder portion (54). 18. The variable compression apparatus for an internal combustion engine according to claim 17, wherein passages connected to the respective spring excrement chambers are configured as orifice passages. A connecting rod 6 which is supported by the piston 8 which reciprocates in the cylinder of the internal combustion engine and which is supported by the crankshaft 1 at the other end. An eccentric sleeve 28 which eccentrically aligns the bearing hole and the piston pin 7 through which the bearing hole is inserted into the connecting rod 6 is provided in the claw support part at one end thereof so as to be rotatable. An eccentric sleeve lock means 29 is provided which can fix the rotation of the eve 28 in the required position, so that the eccentric sleeve lock means 29 is provided from the inside of the connecting rod 6. Moving toward the eccentric sleeve 28 while moving in the direction intersecting with the axial direction of Member 30 protrudes to eccentric sleeve 28 Each surface has first and second engagement portions 28a and 28b which can engage with the pin member 30, and at the same time, the piston portion 30a connected to the pin member 30 is eccentrically sleeved. Of the piston portion 30a so that the pin member 30 can be driven in and out of the connecting rod 6 toward the eccentric sleeve while moving in the direction intersecting the axial direction of the 28. Piston-type fluid pressure drive mechanism (35) having means for generating a differential pressure between both fluid pressure chambers (31) and (32) while the fluid pressure required for the fluid pressure chambers (31) and (32) respectively provided on both sides is applied. Variable compression device of an internal combustion engine, characterized in that provided with. The method according to claim 19, wherein when the pin member 30 engages with the first engagement portion 28a, the pin member 30 is in a high compression ratio state or a low compression ratio state, and the pin member 30 is connected to the second engagement portion ( 28b), the variable compression apparatus for the internal combustion engine, characterized in that a low compression ratio state or a high compression ratio state. 21. The variable compression apparatus for an internal combustion engine according to claim 20, wherein the high compression ratio is set in an engine low medium speed region and a low compression ratio is set in an engine high speed region. The connecting rod (6) is provided with a connecting rod (6) which is supported by the piston (8) reciprocating in the cylinder of the internal combustion engine and the other end by the support of the crankshaft (1). An eccentric sleeve 5 which eccentrically aligns the bearing hole of the connecting rod 6 and the supporting shaft through which the bearing hole is inserted at either end of the support portion at both ends thereof is rotatably provided. By operating the pin member 12 which can move in the axial direction of (5), this pin member 12 is engaged with the engaging portions 5a, 5b formed from the eccentric sleeve 5, A variable compression device for an internal combustion engine, characterized in that an eccentric sleeve lock means (11) is provided which can fix the rotation of the eve (5) in a required position.

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